Power transmission mechanism



J. KNOWLES ET AL POWERy TRANSMISSION MECHANISM April 2, 1963 Filed July21, 1958 5 Sheets-Sheet 1 April 2, 1963 J. KNowLEs ETAL --POWERTRANSMISSION MECHANISM 5 eats-Sheet 2 Filed July 21, 1958 .msm w mkm@ Mwww? m M SSM f A EMM MOM mwN April 2 1953 J. KNowLEs ETAL 3,083,589

POWER TRANSMISSION MECHANISM Filed July 21, 1958 5 Sheets-Sheet 3Tarn/effe 7' )7e ve rse Se rra der 47 l/dfve awr/Sevran a7 d VeINVENToRs %55$2%0- By NORMAN 7." GENERAL April 2 1953 J. KNowLEs ETAL3,083,589

v POWER TRANSMISSION MECHANISM 4&4 ,3194

37a eier f INVENToRs JAMES KNOWLES E E THOMAS R $70CK70N BY NORMAN 7.GENERAL 1C ATTO EVS April 2, 1963 J. KNowLEs ETAL 3,083,589

POWER TRANSMISSION MECHANISM Filed July 2l, 1958 5 Sheets-Sheet 5INVENTORS JAMES mon/Es 32m/ 355/2222 BY 5L@ am ATTORNEYS United StatesPatent O Eames Knowles, Bh'rringham, Thomm R. Slechte-n,

Northvilie, and Norman 'i'. General, Detroit, Micha,

assignors to Ford Motor Company, Dearborn, Mich.,

a corporation of Belawar Filed Early 2l, i953, Ser. No. 74%,725 2i'Claims. (Cl. i4-677) Gur invention relates generally to a powertransmission mechanism and more particularly to a hydrokinetictransmission particularly adapted for use with an engine driven vehiclealthough we contemplate that it may also have a variety of other uses.

yThe embodiment of our invention herein disclosed includes a multipleturbine hydrolrinetic unit and a simple planetary gear unit acting incooperation therewith in a novel manner to provide plural torquedelivery paths from a driving Shaft, such as a vehicle enginecrankshaft, to a driven power output shaft or tailshaft.

The hydroln'netic unit and the gear unit are capable of jointlyproviding a relatively high over-all torque multiplication ratio duringinitial operation from a stalled condition and the magnitude of thisratio decreases uniformly in accordance with the driving torque demandsthroughout a wide range of speed ratios. The hydrolrinetic unit includesprimary and secondary turbine members and a pump member disposed influid ow relationship in a toroidal fluid circuit. 'he portion of theeffective torque output of the hydrokinetic portion of the assembly thatis contributed by the primary turbine member is greater than the torquecontributed by the secondary turbine member during operation in thelower speed ratio range. But and as the speed ratio increases, the ratioof the torque transmitted by the primary turbine member to the torquecontributed by the secondary turbine member progressively decreases.

An overrunning clutch means is provided for establishing a drivingconnection between the secondary turbine member and the primary turbinemember under normal forward driving conditions and for accommodating arelative overrunning motion therebetween during reverse drive and duringforward drive operation in the lower speed ratio range. The primaryturbine member is directly coupled to one element of the planetary gearunit and a brake mechanism is employed for anchoring another element ofthe gear unit to provide the necessary reaction, the remaining planetaryelement being connected to the power output portions of the mechanism.The combined torque of the turbines is thus transmitted to a commonelement of the planetary gear unit after the one-way clutch becomeseffective to establish a torque delivery path between the secondaryturbine member and the primary turbine member, and this combined torqueis multiplied by the gear unit to establish an increased output torque.

After the speed ratio of the hydrolrinetic unit increases, the anchoredelement of the gear unit is released and is synchronously clutched tothe secondary turbine member, a suitable clutch assembly being providedfor this purpose. The two turbine members thereafter operate as a singleturbine member, the effective torque ratio of the gear unit thereafterbeing uni-ty.

it is thus apparent that the transmission assembly is capable ofoperating in two drive ranges; namely, a low drive range in which thegear unit is eective to increase the net output torque of the assemblyand a high drive range in which the torque ratio of the gear unit isunity.

According to a principal feature of our invention, provision is made forobtaining a reverse torque delivery path by using the same gear elementsand converter 3,083,589 Patented pr. 2, 1963 components that areemployed during forward drive. This is accomplished by forming thesecondary turbine member with separate inlet and outlet sections, theoutlet section being adjustable relative to the inlet section to providedifferent effective "blade exit angles for reverse and for forward driveconditions. The exit section may normally assume an angle which isfavorable to forward driving torque delivery, and servo means areprovided for adjusting the effective turbine exit angle to provide areverse turbine torque which in turn is transmitted to the planetarygear elements. -An independent brake band is provided for anchoring theprimary turbine member and one element of the gear unit during reversedrive operation, and after reverse operation is completed the reversebrake may be released and the secondary turbine exit section may beIeadjusted so that the effective secondary turbine blade exit angle isfavorable for forward torque delivery. yThe need for providing aseparate reverse gear mechanism is thus eliminated.

Similarly, the exit section of the secondary turbine member may beadjusted during forward drive operation to the reverse torque positionto obtain a zero torque shift if this is desired so that the sequentialoperation of the low speed brake and the direct Idrive clutch will becharacterized 4by a maximum degree of smoothness.

According to another feature of our invention, provision is made forconditioning the transmission for rela-- tively high -torque ratiooperation to effect maximum acceleration. This is accomplished byforming the pump member with an adjustable blade exit section which maybe adjusted to an angle which is favorable to high torque, high torusflow operation. Suitable servo means are provided within the automaticcontrol system for the transmission yfor effecting an appropriateadjustment of the pump blade exit section. According to a preferredarrangement, the transmission `control system further includes means forconditioning the planetary gear unit for torque multiplication whenmaximum acceleration is desired following normal, cruising operation,and this occurs concurrently with the above mentioned adjustment of the'blade exit section of the pump member to the high accelerationposition.

The provision of an improved hydrokinetic transmission mechanism of thetype set forth above being an object of our invention, it is a furtherobject to provide a transmission mechanism with a multiple turbinehydrokinetic unit and a gear unit coacting therewith whereby theindividual turbine elements of the hydrokinetc unit are effective totransmit a driving torque to a common portion of the gear unit, thelatter being effective to change the magnitude of the effective torquesupplied by the hydrokinetic unit.

It is a further object of our invention to provide a transmissionmechanism having a hydrokinetic unit with primary and secondary turbinesand a gear unit coacting with the turbines and wherein clutch means areprovided for transmitting the torque of the secondary turbine to theprimary turbine, the combined torque output of the primary and secondaryturbines being transferred to a torque input element of the gear unit.

It is a further object of our invention to provide a transmissionmechanism as set forth in the preceding object wherein one element ofthe gear unit may be anchored to provide the necessary torque reactionand wherein another element thereof is directly connected to drivenportions of the mechanism.

It is a further object of our invention to provide a transmissionmechanism embodying a novel arrangement of a multiple turbinehydrolrinetic unit and a gear unit wherein the effective torque outputof the hydrokinetic unit is transferred to a planetary gear unit andwherein clutch and brake means are provided for controlling thetransmission mechanism incorporating a multipley turbineV hydrokineti'cunit and a simple gear unit acting in cooperation therewith wherein thecomponents used for obtaining a forward driving torque delivery may alsobe adapted for delivering a reverse driving torque to the n'ielanisnrthereby eliminating the need for providing additional components forreverse torque delivery.

It is4v a further object of our invention to provide a hydrokinetictransmission mechanism of 'simplified constriiction in which theperformance is characterized by a high degree of smoothness.

It is a further object of our invention to provide a hydrokinetictransmission mechanism of the type abovel set forth in which means areprovided for conditioning the mechanism for hill braking during whichthe torque is transmitted ina reverse direction through the mechanismand readily absorbed by the engine.

Further objects and characteristics of our transmisi sion mechanism willreadily become apparent from the following description of a preferredembodiment of our invention and from the accompanying drawings wherein:

FIGURE 1 is a longitudinal cross sectional view of au assembly view of apreferred embodiment of our invent-ion;

FIGURE 2 is a chart showing the vector diagram for the fluid flow in thehydrokinetic torque converter circuit;

FIGURESv 3a and 3b show a schematic representation of the automaticcontrols of a transmission mechanism;

FIGURE 4 is a cross sectional subassembly view of the secondary turbinemember of the hydrokinetic unit;

FIGURE 5 is a view showing an adjustable blade element of the secondaryturbine member;

FIGURE 6 is an illustration, partly in section, of the secondary turbinemember as viewed in the direction of* the transmission axis.

'For the purpose of particularly describing a preferred embodiment ofthe invention, reference will first be made to the assembly viewillustrated in FIGURE 1 wherein `numeral 10 designates an enginecrankshaft and numeral 12V designates a power output shaft or tailshaft.The

f transmission mechanism usedv for providing multiple torque deliverypaths between the crankshaft 10 and the tailsha-ft 12 comprises ahydrokinetic unit generally shown at 14 and a planetary vgear unitgenerally shown at 16. I'h'e Vhydrokin'etic unit 14 and the gear -unit16 are enclosed by'a housing 18 which may be connected in the usualvfashion to the engine block of a vehicle engine. By preference thehousing 18 is constructed in one piece and, vas Vwill subsequentlybecome apparent, it may be formed withv reducedV dimensions by reason ofthe unique physical arrangement of the elements of the mechanism.

'The crankshaft 19 has secured thereto a drive plate 20 having a starterring gear 22 joined to the periphery thereof in the usual fashion.V YThedrive plate is secured 'by bolts 24 to cooperating flanges 26 andV 28respectively formed on a pump housing 3G and on a housing plate 32, thehousing 30 and plate 32 forming a part of the hydrokinetic u'nit 14. Thepump housing 3G deiines in part a and inner and outer shrouds shown at36 and 38. A primary turbinel member is shown at40 and it includesturbine blades situated at angularly spaced locations about the axis ofthe transmission between an outer shroud 42 and anv inner shroud 44, theentrance section for the primary 'turbine 40 being disposed injuxtaposed relation-l shipv with respect to the exit section of pump 34.A secondary turbine vmember is shown at 46 and it also includes aplurality of blades disposed in angularly spaced torque Yconverted pump34 having radially disposed blades n relationship between inner andouter shrouds 48 and 50, respectively. It is carried by a hub 52 towhich the radially inward portion of the shroud 48 is joined. The' hub52 is positively splined or otherwise secured to a sleeve 54 extendinglongitudinally as indicated.

Hydrokinetic unit 14 further includes a reactor 56 having radiallydisposed blades positioned between the exit section of secondary turbine46 and the entrance sectionl ofthe pump 34, and it functions to alterthe effective fluid entrance angle for the pump 34 in the usual fashionto produce a multiplication of the effective turbine torque. For thisreason the hydrokinetic unit 14 will hereinafter be referred to as atorque converter unit for purposes of the present description.

The reaction S6 includes a hub 58 supported on a oneway brake 6i) whichin turn forms an overrunning connection between the reactor 56 and astationary sleeve shaft 62 which extends axially and which is fixed toan adapted plate 64. Suitable bolts 66 are provided for connecting'plate 64 to a transverse partition wall 68, said wall forming aphysical separation between the torque converter unit 14 and a planetarygear portionY of the mechanism. The wall 68 may be secured to acooperating shoulder on the housing 18 by suitable Ybolts 70 and it isrecessed as shown at 72 to provide a pump chamber for a positivedisplacement gear pump 74. This pump 74 forms a portion of an automaticcontrol circuit for controlling the operation of the various clutch andbrake elements of the transmission mechanism herein described.

The pump housing 30 is permanently secured to a hub shaft 76 which inturn is journaled by a suitable bushing in the wall 68. The shaft 76 isdrivably keyed or otherwise connected to a power input gear of the gearpump 74.

The inner shroud 44 of the primary turbine member 40 is drivablyconnected to a torque transfer disc 7S extending through the radiallyinward region of the torque converter circuit and between the exitsection of the secondary turbine 46 and the entrance section of thereactor 56. The outer periphery of disc 78 is preferably formed with anelongated splined portion as indicated to form a driving connection withshroud 44 and to facilitate assembly. rI'he body of the disc '78 issuitably apertured to permit free ow of iiuid throughout the toruscircuit.

of the converter-'unit without interruption.

The lnner portion of the disc 7S is positively splined to a sleeve shaftdi) which may be journaled by suitable bushings within the stationarysleeve shaft 62.

As previously indicated, sleeve shaft 54 is splined to the hub 52 ofthesecondary turbine 46 and it is also connected to another sleeve shaft 82suitably journaled within sleeve shaft as indicated. The hub 52'of thesecondary turbine 46 also denes an outer race for an overrunning clutchindicated at 84, said clutch forminga one-way driving connection betweenan inner clutch race 8f6'and a centrally disposed shaft SS, the latterbeing splined to the inner race 86. The outer shroud 42 of the primaryturbine member 4# is positively connected to the inner clutch race 86 bymeans of a drive plate 96; The shaft 88 is end supported by suitablebushings within sleeve shaft 82 and within the bearing recess formed ina portion 916 extending through the converter torus circuit in aradially inward direction thereby forming a counection between the innerpump shroud itself and an annular sealing ring identified at 98. Thesealing ring 9SY is received within a circular recess formed in pumpshroud 38 and it defines therewith a radialn passage 10d whichcommunicates with a groove 102,V formed in stator shaft spaanse 62through ports in hub shaft 76. Suitable internal passages are providedfor introducing luid pressure into groove 192 when appropriate, saidpassages forming a part of the automatic control circuit shown inFIGURES 3a md 5b. The portion 96 has formed therein a passage 164extending from radial passage lili? to the annular cylinder 92 and isadapted to distribute working pressure to the piston 94. The blades forpump 34 may be cast integrally with the shroud 36 and the shroud 38 maybe formed with suitable depressions as shown which cooperate withperipheral projections on the pump blades to secure the blades and theinner shroud in place.

rl"he exit section of the pump 34 includes a plurality of blade elements1%6 carried by mounting pins 168 which are mounted for rotation about agenerally radial axis. One blade element may be situated within each ofthe fluid flow passages detned by the generally radially situated pumpblades for the pump 34. It is therefore apparent that when the bladeelements 186 are adjusted from one angular position to another, theeffective pump blade exit angles will be altered. This adjustment of theblade elements 1de is accomplished by means of a mechanical connectionbetween the piston 94 and the pins said mechanical connection comprisingan offset pin portion suitably connected as indicated to the annularpiston 94. When the piston 94 is moved in a longitudinal direction, thepins 162; and the associated blade elements 1126 will be annularlyadjusted around their respective axes.

As previously mentioned, the exit section of the secondary turbine 46 isadjustable to provide variations in the effective blade exit angle ofthe secondary turbine and this adjustment is accomplished by means or" aservo mechanism including a servo cylinder member 11@ hav- 'ng a radialange 112 which in turn is bolted to the rub 52 of the secondary turbinemember, suitable bolts L14 being provided for this purpose. A servocylinder closure wall is formed by a disc 116 which is carried by theaforementioned sleeve 54. A suitable disc washer may be provided asindicated between the disc 116 and the torque transfer disc 73. 'hecylinder member 110 and the sleeve 54 cooperate to define an annularworking space within which an annular piston 11S is situated, the pistoneing adapted to be adjusted in a longitudinal direction. A workingchamber is defined by the piston 118 and the closure disc 116 and theduid pressure may be admitted thereto through a port 129 formed insleeve 54 and the communicating port 122 formed in shaft SS, said portforming a portion of the automatic control circuit for the transmission.The inner periphery of the outer shroud 48 for the secondary turbine 46forms a radial flange and it is disposed between the hub 52 and thecylinder member 11%.

A plurality of radial pins 124 is situated adjacent the exit section ofthe secondary turbine member and journaled in the inner shroud Sil andin cooperating openings formed in the cylinder member 110. Each pin 124carries a blade element 126 and these blade elements are disposedrelatively close to the reactor member 56. The angular position of thepins 124 and the blade elements 12:5 may be adjusted about radial axesand this is accomplished by means of a mechanical connection betweenpiston 11S and an otlset portion 12S located at the radially inward endof pins 124. The offset portions 12S cooperate with a circular grooveformed in the piston 118. This groove is preferably defined by a recess13u and by cooperating retainer ring 132 held in place by a suitable ssnap ring as indicated. t is thus apparent that when fluid pressure isadmitted into the working chamber between the piston 113 and the closuredisc 116, the piston 118 will be moved axially thereby rotatablyadjusting the pins 124 to alter the eective secondary turbine exitangle. It is contemplated that a torque reversal may be obtained whenthe magnitude of the torus ilow velocity is within a predeterminedoperating range and When the blades .125 are adjusted to a reversetorque position and this reverse torque may be utilized, as willsubsequently be explained, to obtain a reverse drive. It is furthercontemplated that when the blades 126 are adjusted to the other extremeposition, the secondary turbine member 46 will be conditioned for apositive torque delivery for effecting forward drive operation.

Sleeve shaft Sil has formed thereon a radially extending frange 134which has secured thereto a brake drum 136. A reverse brake band 138encircles drum 136 and it may be applied by a fluid pressure operatedservo in a conventional manner to anchor sleeve shaft S9.

The shaft 8S is drivably connected to a ring ygear 140 of the planetarygear unit 16 by means of a drive member 142. A planetary sun gear isshown 'at 144 and it is situated in driving engagement with planet gearssuitably journaled on the carrier identified by numeral 14S. The carrier148 is in turn secured to =a power output shaft y for accommodating the:transmission of driving torque from the carrier to driven portions ofthe mechanism.

A brake drum is shown at 152 and it includes Ian inwardly disposed axialextension 154 which may be rotatably journaled on a telescopicallyassociated extension of a supporting member 156 which in turn may besecured by bolts 158 to the rearward portion of the transmission housing13. Extension 154 has secured thereto the sun gear 144 and additionalsupport may 'be provided yfor the sun gear 144 and extension 154 byjournaling the Same on the power output shaft 156.

A friction brake band 16d Surrounds the brake `drumV 152 and it is'adapted to be energized by means of a Illuid pressure operated servo byanchoring the brake drum 152 and the associated sun gear y144. The drum152 and e"- tension 154 cooperate to define an annular working chamber162 within which is slidably situated ian annular piston 164, saidpiston and cylinder cooperating to denne -a iluid pressure chamber. Apiston return spring is shown tat 166 and it is Aadapted to act on thepiston 154, a suitable spring seat 16S being secured to extension 154 bya suitable snap ring to provide a seat for the spring 165. A multipledisc clutch assembly is sho-Wn at 176 and it is eective to provide adriving `connection between brake drum 152 and clutch member 172, thelatter being externally splined to provide a driving connection betweenalternate discs of the clutch assembly 171B` and the brake drum 152being internally splined for accommodati-ng the other discs of theassembly 170. The clutch member 172 is connected .to torque transfermember 174 which in turn is drivably connected to la drive member 176,the latter in turn being secured to sleeve shaft yS2. It is thusapparent that the clutch assembly 179 may be energized by the piston 164as the associated Working chamber is pressurized to provide a drivingconnection between sleeve shaft 82 and sun gear 144.

A positive displacement pump is generally indicated by numeral 178 andit comprises a gear member 189 drivably coupled to power output shaft150 and situated Within a pump chamber delined in part by closure plate182 secured to a cooperating shoulder `formed on the housing 18 by bolts158. The pump 17S, together with the previously described front pumpshown at 72 and 74, forms'a porztion of the automatic control circuitfor the basic transmission structure.

A tailshaft housing 184 is secured to housing 18 by bolts 186 yand it isadapted to enclose a tailshaft friction brake generally designated bynumeral 138. The brake 1S3 includes a relatively stationary casing 1%*secured to the housing 18 -by bolts 192 and a multiple disc brakeassembly `is provided as shown at 194 for anchoring the shaft 1511 tothe casing 19t), said shaft and casing being appropriately splined toaccommodate alternately spaced discs of the assembly 194. Casing 190defines an annular cylinder within which an annular piston is situated,

disc assembly.

Q Referring nex to FIGURES and 6, the blade elements 126 of thesecondary turbine exit section are secured in a suitable fashion, suchas =by projection welding, to a flat 196 formed on pins 124 Aat alocation intermediate the radially inward and outward ends thereof, theradially outward end of pins 124 being received through cooperatingopenings formed in the inner shroud 50 to provide a support forl `thesame. One pin 124 may 'be situated in substantialalignment with .theexit of `each of the blades of the primary section of the secondaryturbine 46. The :bladeV elements 126 may be adjusted as illustrated inFIGURE 5 from -a liirst position which is designated by dotted lines toa second position which is designatedv by full iines.

The blade elements 126'den`ne in part an extension of the fluid iiowchannels in the secondary turbine assembly and they function to adjustthe effective blade exit angle approximately 90'. If it is assumed thatthe secondary turbine Iassembly is rotating in the direction of thearrow lin FIGURE 5, the` direction of the exit uid velocity vector willbe changed so that it will have a forwardly directed. tangentialcomponent rather than atangential'component which extends in a backwarddirection. As will Vsubsequently be explained in more particular detailwith reference to the Yvector diagrams of FIGURE 2, this reversal in theexit iiuid velocity vector will result in a reversal Operation of theTransmission Assembly of FIGURE 1' The transmission mechanism of theinstant disclosure yis capable of providing a relatively high over-allstarting torque ratio and aV -smooth transition from this initialtorqueratio to an over-al1 normal cruising ratio of unity. The planetary gearunit 16 is capable of providing either of two gear ratios during anoperating shift sequence.

Driving power is delivered from the engine crankshaft to the torqueconverter pump 34 and toroidal'uid circulation isthereby establishedwhich causes driving torque to be imparted to .primary turbine member40. If the transmission is operated from a standing start, the magnitudeof the torque acting on primary turbine member 49 is of a substantialdegree lrelative to the pump torque and the direction of the t'orus flowvectors is such that the primary turbine 40 tends to rotate inthe same`direction as the pump 34. However, the blade angularity of the primaryturbine 40 and the secondary turbine 46 is such' that the toroid'al 'uidcirculation establishes a negative moment of momentum in the region ofthe secondary turlbine 46 under these initial star-ting conditions andthe4 Y secondary turbine '46 will therefore tend to rotate in adirectionopposite to the direction of rotation of the pumpY G4' and primaryturbine 40. The uid flow relationship of the'primary `and secondaryturbines is such that the reverse driving torque of the secondaryturbine will prevail until the over-al1 speed ratio reaches a value ofapproximately 0.3.

During these starting conditions, the Vfriction brake band 160 and thedisc clutch vassembly 170 are both simultaneously energized and thiscauses-the sun gear 144 to be anchored so that it may serve las areaction member for the-planetary gear unit 16. The multiplerdisc clutchassembly 170 serves to anchor the secondary turbine 46 since it partlydenes a connection between the energized brake Aband 160 land secondaryturbine, the other portions Vof the torque delivery path for thereaction torque of the 30 as a reaction member.

secondary turbine being dened by torque transfer member 174, drivemember 176, sleeve shaft 82 and hub 52.

The overrunning clutch 84 is ineffective to anchor secondary turbine 46.against reverse rotation and it is 5 thus necessary to utilize themulti-ple disc clutch 170 for 10 The secondary turbine Ifunctionssomewhat as a stationary torque converter reactor during this phase ofthe operation.

I It is contemplated that the automatic control circuit' will eiect -adisengagement of multiple disc clutch assembly 170 after a converterspeed ratio of approximately 0.3 is obtained and the torque actingV onthe secondary turbine 46 will thereafter-be in a forward direction. Theterm converter speed ratio may be defined for present purposes yas theratio of the ring gear speed to the engine' crankshaft speed. Theoverrunning clutch 84 is therefore eifective to transfer the secondaryturbine torque to shaft l8S so ythat the primary turbine torque will besupplemented by the torque of the secondary turbine 46. Both the primaryturbine and the secondary turbine are connected to the common shaft 88and the combined torques are then transferred through shaft 88 into ringgear 140 of the planetary gear -unit 16. The friction brake band 160 iscontinually applied during this stage of the operation and the sun gear144 therefore continues to serve This stage of the operation willAhereinafter :be referred to as the intermediate speed range.

To eifect high speed operation, the multiple disc clutch .assembly 17dmay again be applied in synchronism with the disengagement of the brakeband 16d and the planetary gear unit 16 therefore becomes locked-up toprovide ya gearratio of unity. When a hydrokinetic coupling condition isreached, the reactor 56 overruns in a forward direction by reason of thepositive torque acting thereon,

the one-way brake 6i) permitting such an overrunning motion.

To obtain -high performance operation of the hydrokinetic portion of themechanism, the servo defined by the annular cylinder 92 and piston 94may be -actuated to adjustably position the blade elements 106 at theexit section of the blade elements of the pump 34. This produces Iahigher torque ratio in the hydrokinetic portion of the mechanismalthough the eciency is necessarily reduced.

To obtain hill braking duringcoasting operation, the

brake band 16) may be energized to provide a reaction for the sun gear144 and this causes the ring gear 140 to overspin the power outputshaft. This in turn causes the primary Vturbine 40 to overs'pin. Thereis, therefore, a desired amount of hydrokinetic braking to provide thenecessary deceleration of the vehicle. The multiple disc brake assembly194 may be energized under these con-- ditions by pressurizing theannular working chamber behind the piston 195. This mechanical brakingaction may be controlled by the automatic control circuit in a mannerwhich will subsequently be explained. Provision iS madeV during theoperation of the brake assembly 194 for supplying the discs with asucient amount of cooling oil to prevent overheating.

To obtain reverse drive operation, reverse brake band 138 may beenergized and the friction brake band 160l may be de-energized. Also,thermultiple disc clutch aS- sembly 170 is energized to establish aconnection between the sun gear 144 and the torque transfer member 174,-the latter being connected to the secondary turbine member 45.- `I-naddition, the working chamber defined by the annular piston 118 and thecylinder member 110 iS pressurized to eiect a shifting movement of theblade elements 126 of the secondary turbine member 46. This,V

as previously mentioned, produces a reverse torque on the secondaryturbine member 46 and this reverse torque is transferred through sleeveshaft 82 and through the energized multiple disc clutch assembly E70 tothe sun gear 144. The secondary turbine member therefore drives the sungear M4 in a reverse direction, and since the reverse brake band 13S isenergized, as previously explained, the carrier 148 and the associatedpower output shaft 150 are driven in a reverse direction.

Referring next to the vector representations or" FIG- URE 2, the luidtlow relationship between the various components of the hydrokiueticportion or" the mechanism is illustrated vectorially. The vectordiagrams of FGURE 2 represent the motion of a particle of fluid in thehydrolsinetic torus circuit at the exit of each of the convertercomponents. A separate diagram is shown for the stalled condition andfor a speed ratio of 0.5 under forward driving conditions. Similarly,vector diagrams are provided for the stalled condition during reversedrive and at a speed ratio of 0.5 during reverse drive. The tangentialvelocity vector of a point on the various blade elements is representedby the symbol U and the duid low velocity component along the blade isrepresented by the symbol W, the latter being at an angle equal to theangle of the blade elements. The resulting absolute velocity ector isrepresented in each instance by the symbol V and the vector F representsthe fluid flow through the torus circuit in a direction normal to theplane of rotation of the blade elements. The angles ot the inlet andexit sections of the various blade elements in the circuit are indicatedand this is the preferred geometry although it is possible thatdeviations may be made therefrom depending upon the performancecharacteristics which are desired. ln the case of the stator exitvectors, the vector component U is absent since the stator is anchoredagainst rotation during the converter phase of the operation and theabsolute velocity vector is equal to and coincident with the vectorrepresenting the ow along the blade. rthe torus flow velocity dccreasesfrom a maximum at stall to a lesser value at 0.5 speed ratio in bothforward and reverse driving ranges and this in turn results in acorresponding decrease in the absolute fluid ilow velocity although thedirection of the fluid velocity vector remains unchanged.

in the vector diagram for the pump or impeller, the absolute iloivvelocity vector W is the resultant of the vectors U and V. Duringforward drive the torus iiow velocity vector decreases from a maximum atstall to a lesser value at 0.5 speed ratio and a similar decrease occursduring reverse drive, the eect of this change on the magnitude of theabsolute iiow velocity being relatively minor. The exit section or" theimpeller blade elements may be Ishifted to the performance positiondesignated in FGURE 2 and this alters the angle of the vector W.However, this change in the direction of vector W has only a minoreffect on the vector V since the vector U is substantially increased dueto the resulting increase in the pump speed.

The vector diagram for the first turbine exit illustrates the erle-ct ofchanges in speed ratio upon the absolute velocity vector V as the speedratio changes. lt may be observed that as the speed ratio changes from astalled condition to 0.5 speed ratio during forward drive, the vector Vchanges from a backward direction to a forward direction; and as theangularity of the vector V approaches the angularity of thecorresponding vector for the impeller exit, the primary turbine torqueapproaches zero. This is true since the moment ot momentum of the iluidwhich enters the iirst turbine is equal to the moment of momentum ot thefluid leaving the impeller lades and if the change in the moment ofmomentum for a particle of i'luid passing through the first turbineapproaches Zero, the turbine torque will also approach zero. Forpurposes of discussion, the moment of momentum of a particle of fluid inthe torus circuit can be defined as the mass of that particle multipliedby the tangential component or" the absolute velocity vector V mull0tiplied by the radial distance of that particle of fluid from the axisof rotation.

During reverse drive the primary turbine is anchored by reverse brakeband 13S and, accordingly, the vector U is lacking in the vectordiagrams for reverse operation. lt is thus seen that the only change inthe absolute velocity vector for the iirst turbine during reverse driveas the speed ratio changes from zero to 0.5 is due to the change in. themagnitude of the torus ilow velocity.

Referring next to the vector diagrams for the second turbine exit, it isseen that the angularity of the absolute velocity vector changes ratherrapidly as the speed ratio changes. This can be observed by comparingthe vector diagram for stall with the vector diagram for 0.5 speed ratioduring forward drive wherein the vector V changes from au angle ofapproximately to an augle of approximtaely 95.

rthe moment of momentum of a particle of iiuid leaving the first turbineis equal to the moment of momentum of a particle of fluid entering thesecond turbine and he change in these quantities is a measure of thetorque cting on the second turbine. lt should be noted that the absolutevelocity vector for a stalled condition for forward drive range issubstantially equal in magnitude and dhection as tac correspondingvector for the rst turbine. lowever, since the exit section of thesecond lrbine is situated at a radius which is smaller than theoperating radius of the second turbine entrance, a change in the momentof momentum of a particle of iluid passing through the secondary turbinefor a stalled condition in the forward drive range is accomplished. Itis thus apparent that torque will be imparted to the second turbine in areverse direction and the second turbine will thus tend to turn in areverse direction when the converter is conditioned for forward driveand when it is operated at speed ratios near zero. However, aspreviously eX- plained, the secondary turbine is anchored by thesi1nulteneous engagement of the multiple disc clutch N0 and the frictionbrake band to pre 'ent such reverse r0- tation.

It should be further noted that when a speed ratio of 0.5 is obtainedduring lforward drive operation the absolute velocity vector for thefirst turbine exit and the second turbine entrance will have experienceda substantial change in angularity and the tangential component or .thevelocity vector will have changed from a reverse direction to a forwarddirection. It is this change in angularity which results in a change inthe `direction in Wluch the moment of momentum operates. As previouslyexplained, the multiple disc clutch assembly is de-energized to permitthis positive torque of the transmission to be transferred through theoverrunniug clutch Se so that it will be added to the torque contributedby the iirst turbine.

Referring next to the eXit velocity vectors for the second turbineduring reverse drive, it is apparent that the angularity of vector W hasbeen changed from an angle greater than 90 to an .angle less than 90 byreason of the adjustment ot the secondary turbine exit blade elements126 yby the associated servo mechanism described in connection withFlGURES 1 and 5. Referring more particularly to the vector representingthe stalled condition for reverse drive range, the tangential componentof the absolute velocity vector at the primary turbine exit or thesecondary turbine entrance has a negative sense so that the change inthe moment of momentum of a particle of tluid passing through the secondturbine will be very substantial. Since it is negative in character, thesecondary turbine tends to rotate in a reverse direction. However, asthe speed ratio increases during reverse drive, the angularity of thekabsolute velocity vector for the secondary turbine exit changes rapidlyso that it tends to become aligned with the absolute velocity vector forthe primary turbine exit or the secondary turbine entrance. Theresulting decrease in the change in moment Description of the AutomaticControl Circuit of FIGURES 3a.' yl1/2d 3b It is contemplated that theshift sequences previously described may take place automatically inaccordance with the operating road requirements. In the particularcircuit hereindisclosed -a dual driving range feature is incorporated,yalthough other schematic larrangements may also be employed which donot incorporate such a dual driving range feature but which do embodyother characteristics. g

During operation in the rst forward driving range, the multiple discclutch assembly 170 is continuously energized and the friction brakeband '160 is energized during operation from a standing start until adirect drive is accomplished. The control circuit includes an automaticshift valve responsive to engine -throttle setting and to vehicle speedto yde-energize the friction brake band 160 when a high speed cruisingcondition is obtained. This arrangement provides a high degree ofsmoothness during acceleration from a standing start although it is notthe most desirable arrangement from a performance standpoint in certainspeed ratio ranges since the simultaneous actuation of the multiple discclutch assembly 170 and the friction brake band 160 causes the secondaryturbine member 46 to remain stationary. At certain speed ratios in theintermediate range the angularity of the -blade elements Ifor thesecondary turbine member 46 may be unfavorable with respect to theangularity of the absolute duid velocity vectors in the torus circuitwhen theV secondary turbine is stationary. Y

The control mechanism as illustrated in the circuit of FIGURES 3a and 3bis also capable of operating in a second drive range in which themultiple disc clutch 170 andV the friction brake -band 160 may besequentially operated to provide both an intermediate gear ratio and ahigh speed gear ratio of unity. The multiple disc clutch assembly 170and friction brake band 160 are each simultaneously energized betweenzero speed ratio and approximately 0.3 speed ratio, as previouslymentioned, in order toy -obtain the most desirable performance duringthe initial stages of acceleration from a standing start.

The control circuit of FIGURES 3a and 3b includes the aforementionedengine driven pump 74 and tailshaft Vdriven pump 178, both of whichcooperate to provide a source of control pressure for the circuit. Thereverse brake bandlSS is energized-by -a uid pressure operated servo 196which includes a piston 198 acting within a cooperating cylinder, asuitable linkage mechanism 200 being provided for transferring the servoforces to the brake band. The piston 19S is urged to la brake releasedposition by return spring 202.

` A secondfbrake servo is provided at 264 for operating the: low speedfriction brake Aband 160 and it includes a fluid pressure operatedpiston 206 situated within a servo cylinder, the piston and cylinder forthe servo 204 cooperating to define opposed pressure chambers onopposite sides of fthe piston 266. A piston return spring is provided at20S for normally urging the piston to a brake'de-energized position.Suitable force transmitting means are provided as indicated between thepiston 206 and the friction brake band 166.

The fluid pressure operated servo for the multiple disc clutch yassembly170 was previously described in connection with FIGURE 1 and it will bereferred to generally in the subsequent description of the controlcircuit bynumeral 210,

A main regulator valve is shown in FIGURE 3a `at 212 and it includes amultiple land valve spool having spaced valve lands 214, 216 and 218.The region of the main regulator valve between valve lands 214 land 216communicates with control pressure passage 220 which in turn isconnected to control pressure passage 222 extending to thedischarge sideof the tailshaft driven pump 178. Passage 220 also communicates with thedischarge side of the engine driven pump 74 as indicated. kAnotherpassage 224 extends rfrom the discharge side of the engine driven pump'74 to the main regulator valve at a location adjacent valve land 21S,the latter being Iadapted to control the degree of communication betweenpassage 224 and a low pressure exhaust passage 226 communicating with atransmission sump. Similarly, valve land 21,6 controls the degree ofcommunication between passage 22u and the exhaust passage 226.

The regulator valve element is biased in a downward direction yasindicated in FIGURE 3a by a valve spring 22S and Ithis spring force isopposed by a fluid pressure force acting on a valve plunger 239 which isadapted to engage the lower end of the regulator valve element. Valveplunger 230 is pressurized by means of'a branch passage 232.

The discharge side of pump 74 4communicates with` plunger 230 through aone-way check valve 234 and a second one-way check valve 236 is situatedas indicated in FIGURE 3b between lthe discharge side of the tailshaftdriven pump 178 -and passage 222. When the vehicle is operating at lowvehicle speedsV and when the Vehicle is stopped with the engine running,the discharge pressure of the tailshaft driven pump 178 is either zeroor is reduced in magnitude relative -to the discharge pressure `forpumpY 74. It is thus apparent that check valve 236 will be closed underthese conditions and check valve 234 will be open because ofdifferential'pressure. Fluid pressure is thus entirely suppliedwby thepump 74 through passages 221i and 224. The valve land 216 will blockcommunication between passage 220 and exhaust passage 226 under theseconditions and the valve land 21S provides necessary pressure regulationby appropriately controlling the degree of communication between passage224 and exhaust passage 226, the balanced forces acting on the regulatorvalve being provided by spring 228 and spring pressure force on plunger230.

After the vehicle speed begins to increase, the tailshaft driven pumppressure increases relative to thepump discharge pressure of pump 714and under certain conditions it will exceed the discharge pressure forVpump 74.`

For example, during coasting or during a push start the check valve 234will close and check valve 236 will be opened, and the rear tailshaftdriven pump 178 will Vthen form the pressure source for the controlcircuit.

Under these conditions the valve land 216 provides the necessarypressure regulation by Vcontrolling the degree of communication betweenpassage 220 and exhaust passage 226 and this results in a shiftingmovement of the valve spool in an upward direction. The discharge fromvalve lands 246 and 242 communicates directly withY control pressurepassage '2222 by another control pressure passage 244, As schematicallyillustrated, the valve element for the manual valve 238 may be adjustedto either of iive operating positions designated in FIGURE 3a by thesymbols R, N, D, Ds and HB which respectively correspond to a reversedrive range position, a neutral posilg tion, a irst drive rangeposition, a second drive range position and a hill brake position.

'ihe position of the various valve elements illustrated in FlGUlES 3aand 3b correspond to the irst drive range position in which the multipledisc clutch assembly 17? is continuously engaged throughout the shiftsequence as the vehicle accelerates from a standing start to a steadystate cruising condition. The servo 2l@ for the multiple disc clutchassembly 171') is pressurized oy means of a passage 2li-6, a hill holdvalve 24S, a passage 25%, a shuttle valve 252, passage 254i and passage256, the latter communicating with the pressurized portion of the manualvalve chamber. The hill hold valve 248 includes a multiple land valvespool 258 which is spring urged in a left-hand direction as viewed inFIGURE 3b, and when it is in the position shown it provides freecornrnunication between passages 246 and 25d. 'lhe aforementionedshuttle valve 252 includes a single valve plunger 26% and it is urged ina left-hand direction as indicated during operation in the first driverange under the influence of control pressure to permit treecommunication between passages 254 and 254i.

A low servo shift valve is generally identified by numeral 252 and itincludes a multiple land valve spool 264 and a valve plunger 265. Theplunger 266 and the valve spool 264 are situated in separate portions ofa cooperating valve chamber which are separated by a separating element268. A force transmitting stem 27@ is interposed between the plunger 265and the valve element 264 and it is adapted to slide freely in asepmating element 253. The valve 261i includes three spaced valve lands272, 27d and 276 and it is urged in a leit-hand direction by a valvespring as indicated.

During operation in the low speed ratio range while the manual valve isin the lirst drive range position, the valve element 26d. and plunger265 will assume a left-hand position under the influence of the valvespring and communication is thus established between passage 254 andpassage 278, the latter extending to the apply side of the low speedservo 204. An oriiice control valve 28u is situated in passage i273 andpartly deiines passage 273. The orifice control valve 239 includes avalve element having spaced valve lands which is normally urged in adownward direction by an associated valve spring. But when the valveelement assumes the position shown, free communication is establishedtherethrough. When the orifice control valve element assumes a downwardposition, a first branch portion 282 of the passage 270 is blocked bythe upper valve land and a second branch passage portion 234i isuncovered by the lower valve land, the latter branch passage portionhaving a ow restricting orilice 286. It is thus apparent that freecommunication through the oriiice control valve is interrupted and atluid flow restriction is introduced in the passage 270 when the valveelement therefor is moved in a downward direction.

The oriice control valve is subjected to a pressure signal which issensitive to engine throttle movement and this pressure signal, whichwill hereinafter be referred to as throttle pressure, acts on the lowerend of the orifice control valve element and normally biases the same inan upward direction whenever the same is under torque. This samethrottle pressure acts on the differential pressure areas of spacedvalve lands 272 and 274- of the low servo shift valve 262 to urge thesame in a left-hand direction to supplement the action of the low servoshift spring.

The valve actuating forces acting on the low servo shift valve element262 in the left-hand direction are opposed by a vehicle speed sensitivepressure signal acting on the left end of valve plunger 266. This speedsignal ereinafter be referred to as governor pressure and the portion ofthe circuit which produces this governor pressure, as well as theportion of the circuit ld which produces the throttle pressure, willsubsequently be fully explained.

At stall and at lower vehicle speeds, the low servo shift valve elementwill assume a left-hand position and tree communication will beestablished between passage 25dand passage 278. rl`he low speed servowill thus become energized. The transmission is then conditioned foracceleration from a standing start since both the multiple disc clutchassembly i7@ and the low speed servo Zildare simultaneously energized.When the vehicle speed increases, the magnitude of the governor pressureforce will correspondingly increase and at a certain shift point thevalve element 2nd will be urged in a right-hand direction as viewed inFEGURE 3a. The apply side of the low speed servo 284 will then beexhausted through passage 2'le and through an exhaust port in the lowservo shift valve chamber. The multiple disc clutch assembly 17?contirues to be energized following this shift and the transmission willthereafter operate in a cruising, direct drive ratio.

A throttle valve mechanism is generally designated in FIGURE 3a bynumeral 2% and it includes a valve spool 2% having' spaced valve landsthereon and a hollow valve element 292. Another valve element 294 istelescopically received within valve element 292. A valve spring 2% isinterposed between valve elements 292 and 2% and a snap ring 2% isprovided for limiting the amount of relative movement between the twoValve elements. Another' valve spring is interposed between valveelement 294 and valve spool 296. The valve element 292 is connected tothe engine throttle by a suitable mechanical linkage so that the enginethrottle will cause a corresponding movement of the valve element 292.When the engine throttle is advanced, spring 304B becomes compressed andre biasing force acting on valve spring 2% is correspondingly increased.

The valve cham er for valve spool 290 communicates with control pressurepassage 244 by means of passage 3%2 and a throttle pressure passage Silealso communicates with the chamber for the valve spool 296 at a locationintermediate the spaced valve lands. The valve spool Zsll controls thedegree of communication between passages 332 and 3M so that the biasingforces acting in a right-hand direction on the valve spool 29@ will tendto increase the degree of communication between passages 3il2 and S94thereby tending to increase the magnitude of the resulting throttlepressure in passage 36d. Throttle pressure also is caused to act on theright-hand side of valve element 294i and, if desired, a suitablesecondary valve spring may be provided at that location. rThe force ofthe throttle pressure acting on the right side of the spool valve 29hthus balances the biasing force of spring 30d.

The throttle pressure thus produced in passage 34M is conducted to thelow servo shift valve and the orifice control valve as previouslyexplained.

Throttle pressure is also conducted to a main shift valve generallydesignated by numeral 3%. This main shift valve comprises a multipleland valve spool 3% slidably situated in a cooperating valve chamber.The valve spool 368 is biased in a left-hand direction as shown inFIGURE. 3b by valve springs 314i and 3l2. A throttle pressure modulatorvalve element 314 is situated at one end of the shift valve chamber andthe aforementioned valve spring 312 acts directly thereon to urge thesame in a right-hand direction. Throttle pressure is caused to act onthe right side of plunger 314 and when the shifted valve spool 368 is ina left-hand position, the valve element 32M- operates to regulatecommunication between passage 3M- and a modulated throttle pressurepassage Sie, said passage 3l6 communicating with the right-hand side ofthe valve chamber for valve spool Stili within which springs 3l@ md 312are situated. Modulated throttle pressure also is caused to act on adifferential area formed by Valve lands 318 and 329 on the valve spool30S for urging the latter in a left-hand direction. When the manualvalve assumes the rst drive range position as illustrated, controlpressure does not communicate withY the valve chamber for valve spool308 and the main shift valve is inoperative although it may be causedtoshift under the inuence of the opposing forces established by themodulated throttle pressure and by the governor pressure. Governorpressure is caused to act on the left end of valve spool 308 therebyestablishing a pressure force which is proportional to vehicle speed andwhich acts on the shift valve spool in alight-hand direction.

K Governor pressure is obtained in the particular embodiment hereindisclosedby means ofthe tailshaft driven pump 178, although it iscontemplated that other governor pressure sources may also be used. Thepump 17-8 is formed withtWo-working regions, one of which presandthermain shift valve 306, as previously mentioned,

to provide the necessary vehicle speed signal for establishing shiftpoints.

A compensator valve is shown in FIGURE 3a at 33t) and it is capable ofappropriately modifying the operating control pressure level inaccordance with vehicle speed and engine torque. Compensator valve 330includes a valve spool 332 operating in a cooperating valve chamber anda compensator valve spring 334 is situated in the valve chamber forbiasing the valve spool 332 in a righthand direction as viewed in FIGURE3a. Control pressure is distributed to the compensator valve chamberthrough a branch passage 336 at a location intermediate valve lands 33Sand 34% on the compensator valve spool 332. A compensator pressurepassage 342 communicates with the compensator valve chamber at alocation adjacent Valve land 340 and it extends to the main regulatorvalve 212 at a locationadjacent valve land 214. The compensator pressurein passage 342 is capable of exerting an upwardlyV directed force on themain regulator valve spool in opposition to the force of spring 228.Valve land 340 of the compensator valve controls the degree ofcommunication between branch passage 336 and compensatorY passage 342.

Throttle pressure is distributed to the right-hand side of the valvespool 332 through the passage 343-which communicates with passage 304through a low boost valve generally identified by numeral 344. Thisvalve 344 will be subsequently described. Also, governor pressure isdistributed to the compensator valve chamber on theleft side of thevalve spool 332 by means of a Y Vpassage 346 which communicates with thepreviously mentioned governor passage 328 through the low boost valve344. Thefthrottle pressure and governor pressure therefore establishopposed forces which influence the degree of communication betweenpassages 336 and 343 that is provided .by valve land 340. An annularworkingarea is formed on the valve spool 332 at the right-hand side ofvalve land 340 and compensator pressure passage -342 thereforeestablishes a valve biasing force which tends toreduce the degree ofcommunication between passages 336 and 342. The force differentialproduced by the governor pressure and by the throttle pressure, theforce of the spring 228 and the valve biasing force of the compensatorpressure produce a balincrease for any given engine throttle setting,the degree.

Y trol pressure in the circuit which in turn reduces the anced forcesystem and it is thus apparent that theV capacity of the iluid pressureoperated servos to a value which is commensurate with the reduced torquere-V quirernents. lt is pointed out that an increase in vehicle speedfor any xed engine throttle setting is tantamount to an increase in theover-all speed ratio and to a decrease in over-all torque ratio. Therelationship be-v tween speedratio and torque ratio can be readilycalculated. i

Conversely, if the engine throttle setting isincreased for any givenvehicle speed, the compensator pressure will be decreased and thecompensator pressure force acting on the main regulator valve spool willalso bedecreased. This produces an increase in the operating controlpressure Vand the capacity of the transmission servos is correspondinglyincreased to accommodate the increased torque requirements whichaccompany the ad-v justed engine throttle setting.

It is undesirable to allow the compensator pressure to become increasedfor any -given throttle setting after a certain limiting speed isobtained. Otherwise, thel operating control pressure will be reduced athigh cruising speeds to a value which is insutcient to maintain therequired torque capacity of the transmission servos. For this reason agovernor pressure cut-out feature is provided in the compensator valveand .this includes the valvev plug 346 situated adjacent the compensatorvalve spool` 340 and it is provided with spaced valve lands forvdefining a differential Working area 348, said working area beingvin uidcommunication with branch passage 336 so that it is subjected to linepressure. The uid pressure force of the line pressure is opposed by apressure force established by governor pressure, the latter acting onthe right-hand side of valve plug 346. A spacer element. 350 is movablydisposed between valve spool 332 and Valve plug 336.

During operation of the transmission at lower vehicle speeds, thegovernor pressure is insuicient tol establish aV force on valve plug 346which Will overcome the `opposing pressure force of the line pressureacting thereon and the valve plug 346 will be urged in a right-hand di,-rection so that .it isinoperative. However, as vehicle speed increasesfor any given throttlel setting, a point isV reached at which thegovernor pressure will overcome the opposing force of the controlpressure acting on` valve plug 346 and the plug 346 will be urged in aleft-hand,

direction. The force differential` acting on the plug 346 will then betransferred directlyV to theV valve spool 332 through the spacer element350. The area ofv plug 346 upon which governor pressure acts is equal tothe area of the valve spool 332 on which the governor pressure acts andfurtherincreases in governor pressure will create balanced andopposedforces on the valve spool 332. The compensator valve willthereafter be insensitive to changes in vehicle speed and furtherchanges in control pressure below a safe minimum4 value will thereforenot occur.

Let itnow be assumed that the manual valve 23S is' The right hand end ofthe VValve spool for blocker valve 353 is also subjected to linepressure by means of passage 360 which in turn communicates with thepreviously described pressurized passage 254. A blocker valve spring isdisposed on the left side of the blocker valve spool and when passage ispressurized, the valve spring urges the blocker valve spool in aright-hand direction. When this occurs communication is establishedthrough the blocker valve by the governor pressure passage 328 and thepassage 362, the latter extending to the left-hand side of valve spool25S for the hill hold valve 248. During operation in the rst drive rangeposition, the Iblocker valve is maintained in a left-hand position underthe influence of control pressure acting on the right-hand side thereofand the passage 362 is con- :tinuously exhausted ,through an exhaustport associated with the blocker valve chamber.

The hill hold valve is comprised of three spaced valve lands identifiedby numerals 364, 3de and 36S and it is urged in a left-hand direction bya valve spring 370. It is thus apparent that the valve biasing forceestablished by the governor pressure will oppose the spring biasingforce and when the governor pressure exceeds a calculated value, thehill hold valve spool will shift in a right-hand direction so that valveland 354 will block passage 25? and thereby establish communicationbetween passages 246 and the passa-ge 372 between the spaced valve lands36e and 365, Passage 372 in turn extends to the fluid pressure chamberon the release side of the piston 236 of the low speed servo 2174.Passage 372 also communicates with the manual shift valve chamber at alocation between spaced valve lands 374 and 376 on the main shift valvespool.

Passage 362 is also in communication with the hill hold valve chamberadjacent the valve land 3dS and when the hill hold valve spool isshifted in a right-hand direction under 'the influence of governorpressure, the differential area provided by valve lands 366 and 36S issubjected to governor pressure and a snap action occurs whereby the hillhold valve spool lis shifted quickly in a right-hand direction as soonas valve land 368 uncovers the associated branch portion of passage 362.

During operation in the second drive range, control pressure isconducted to the right-hand end of the low servo shift valve spool 264and to the right-hand end of valve plunger 266. Valve spool 264 istherefore shifted in a left-hand direction, thereby establishing .fluidcommunication between passage 254 and passage 278. Valve plunger 266remains inoperative since the governor pressure is insuicient toovercome the opposing force of the line pressure acting on plunger 2bn.It is thus apparent that the low servo shift valve remains in aleft-hand position during operation in the second drive range and isinsensitive to changes in vehicle speed.

If it is now assumed that the vehicle is operated from a standing startwith the manual valve adjusted to the .DS position, the servo 2li) andthe apply side of the brake servo 204 will be simultaneously energizedduring the initial stage of the shift sequence thereby providing formaximum torque multiplication in the hydrokinetic portion of thetransmission mechanism. The release side of the servo 264 is exhaustedthrough passage 372 and through a communicating passage 373 whichextends to an exhaust port located in the main shift valve charnber.After the vehicle begins to accelerate, the lgovernor pressure will urgethe hill hold valve spool in a right-hand direction against the opposingforce of the spring 37) to establish communication between passage 24oand passage 372, and the clutch servo 2id is therefore exhausted throughthe hill hold valve and .through passage 572, passage 373 and the mainshift valve exhaust por-t. Simultaneously with this shifting movement ofthe hill hold valve spool, valve land 264 blocks passage ZStl. Thetransmission is thus conditioned for an intermediate driving torqueratio and both the primary and secondary turbines are operative todeliver torque to the ring -gear 14) of the planetary gear unit 16, thetorque of the primary turbine being transferred to the secondary l@turbine through the overrunning clutch 84 as previously explained. Thelow speed friction brake band i653 continues to anchor the sun gear 1614of the planetary gear unit lo to provide the necessary torque reaction.

When the vehicle continues to accelerate during operation in the seconddrive range for any given engine throttle setting, the governor pressurewill ultimately increase to a value which will be suffi ient to urge themain shift valve spool in a right-hand direction against the opposingforce of valve spring 3l@ and the modulated throttle pressure forcesacting on the main shift valve spool. rhis position of the main shiftvalve` spool is illustrated in FIG- URE 3b, and when it is in thisposition communication is established between passage 354 and passage373 and the main shift valve exhaust port is no longer in communicationwith passage 37S, the release side of the low speed servo thereforebeing pressurized. As previously explained, the appply side of the servo204 is also pressurized, but since the servo spring Zo normally urgesthe piston to a retracted position, the low speed servo becomes releasedunder spring pressure.

Since passage 372 communicates with passage 2li-6 through the11p-shifted hill hold valve element, the passage 246 and the clutchservo 2li? become pressurized following an up-shift of the main shiftvalve element and the planetary gear unit "ad therefore assumes a'locked-up condition. The transmission is thus conditioned for cruising,direct drive operation.

lf hill braking is desired, the manual valve may be shifted to the HBposition and this causes the passage 378 to become uncovered by manualvalve land 242 and the associated exhaust port for passage 378 issimultaneously closed by this same valve land. Passage 37d communicateswith passage 355i which in turn communicates with a passage 3S?. throughthe `throttle valve chamber, a suitable groove 334 being provided in thevalve element 292 for this purpose. Line pressure lin thus conducted tothe left-hand side of the blocker valve generally shown at 38o, therebyshifting the same in a right-hand direction against the opposing forceof a blocker valve spring. The blocker valve thereby blocks passage 352.The passage 372 establishes communication between passage 254 andpassage 352 so that the latter continues to be pressurized with controlpressure. Passage also extends to the main shift valve chamber anddistributes control pressure to the right end of the main shift valvespool and to the differential area provided for the valve lands 313 and32d to shift the main shift valve spool in a lefthand downshiftposition. Communication is thus established between passage 3S@ and apassage 38S which in turn extends to the left-hand side of the low boostvalve 344 to urge the `latter against an opposing spring force to aright-hand position. This interrupts communication between passages 328and 345 and simultaneously exhausts passages Sell and 34S through theassociated exhaust port. In addition, the low boost valve establishescommunication between passage 342 and a converter pressure passage 3%,said passage 39d extending to the interior of the hydrokinetic toruscircuit. When tne low boost valve is shifted in the right-hand directionin this fashion, the governor pressure acting on the left-hand side ofthe compensator valve spool is exhausted and the throttle pressure whichacts on the right-hand side of the compensator valve spool is replacedby a relatively high converter pressure. Both of these pressure changesresult in an in- .reased biasing effect on the compensator valve spoolin a left-hand direction which ytends to decrease the degree ofcommunication between passages 336 and 342, and this in turn results ina decrease compensator pressure. This decreased compensator pressure inturn results in an increased control pressure in the circuit and thecapacity of the brake servo 264- will accordingly be increased. Theincreased line pressure will prevent slippage during a coastingoperation and although the friction brake band 16d is self-energizingfor torque delivery during normal forward driving, the brake capacity isstill sucient during delivery of torque in a reverse direction duringcoast even though the self-energizing feature of the brake-band islacking.

It is thus apparent that the release side of the low speed servo 294will be exhausted through lthe main shift valve exhaust .port and thatthe low speed servo will become applied. This anchors the sun gear ofthe planetary gear unit i6 to cause an overspeeding of the ring gear14)y of the primary turbine member 40.

The blocker valve 358 continues to establish communication betweenpassages 328 and 362 to distribute governor pressure -to the left-handside of the valve spool of hill hold valve 43 and when the `main shiftvalve is downshifted as above described, the clutch servo 21C* isexhausted through the passage 246, through the hill hold valve andthrough passages 372 and 37S, the latter communicating with the mainshift valve exhaust port. As deceleration continues to a point where thegovernor pressure is insuflicient lto maintain the hill hold valve inthe up-shift position, the hill hold valve spool will be shifted to theposition shown in FIGURE 3b and the clutch servo 219 will again beenergized so that both the servos 219 and 294 are simultaneouslyenergized. The secondary turbine member 46 is anchored during this stageof the hill braking optration.

Additional hill braking may be obtained by the multiple disc brakeyassembly 194. .The automatic controls for this brake assembly includesa retarder activator valve 392 which comprises a simple valve spoolbiased in one direction by a valve spring `and in another direction bythrottle pressure which is distributed thereto by a passage 304. Whenthe valve spool for the retarder activator valve 392 assumes theposition shown, communication is established between the aforementionedpassage 254 and a passage 394. However, during a coasting operation thethrottle valve pressure drops to zero and the retarder activator Valvespool assumes a right-hand position, thereby connecting theaforementioned governor pressure passage 328 with passage 394. Passage394 in turn extends to one side of'a retarder clutch valve, identied bynumeral 396, which comprises a simple valve spool normally urged in aleft-hand direction by an associated valve spring. A line pressurepassage 398 communicates with the valve chamber associated with theretarder brake valve spool and when the valve spool is Ipositioned asshown in FIG- URE 3b, the passage 398 is blocked. However, when theretarder brake valve spool assumes a left-hand position, passage 39Scommunicates with a passage 400. During normal driving operation, the:throttle pressure is sufficient to maintain the retarder activatorvalve spool in a left-hand position and the retarder brake valve is thuscontinuously urged in a right-hand position under the influence ofcontrol pressure. However, during coasting position the retarderactivator valve will move under spring pressure in a right-handdirection thereby pressuriz'ing the retarder brake valve spool withgovernor pressure rather than control pressure. At relatively highspeeds the governor pressure is suicient :to maintain the retarder brakevalve in the position shown. However, after the vehicle acceleratesbelow a calculated speed, the retarder brake valve will shift in a lefthand position thereby establishing communication between line pressurepas- Vsage 39S and passage 460, the latter in turn extending to theretarder brake disc lubrication jets 492. When the retarder brake valvespool is shifted in this fashion in a lefthand direction, communicationis established between a passage494 and a passage 466, the latterextending to the Iservo for the brake assembly 194 which is partlydefined by piston 195. Passage 44 is pressurized with the modu- =latedretarder pressure obtained by means of a retarder pressure regulatorvalve identied by numeral 408. Valve 49S comprises a valve spooloperating in an associated valve chamber which communicates with controlpressure passage 254. The retarder regulator valve spool is urged in'aright-hand direction by valve spring 4.1(hand-` the biasing effortof-valvef spring- 410- tends' to increasesY multiple disc retarder brakeassembly 194 wil-l be deter-- mined by the calibration ofthe retarderYregulator valve.

If a forced kickdowm is desired following cruising, direct driveoperation, theY valve element 292 .mayl be urged to the extremeright'position by depressing the accelerator pedal which moves theengine throttle to its extreme position. When this is done -a port `412in the valve element 292 is broughtl into communication with the branchpassage 414 extending from throttle pressure pas.- sage 304. This thenadmits throttle pressure into the interior of valve element 292 on oneside ofthe associated valve element 294, suitable passages 416 beingprovided for this purpose. Continued movement of the valve element 292to a downshift position will cause Contact between a shoulder 418 on thecontrol valve body and an abutment 422 on the valve element 294. As thevalve element 292 moves further, relative movement will then take placebetween valve elements 294 and 292 and-this relative movement isresistedby the throttle pressure which Vexits. in the interior of valveVelement 292. The vehicle operator therefore experiences what may bereferred to as a fluid detent feel which .permitshirn to determinereadily when a downshift will be initiated. This also prevents aninadvertent downshift by the vehicle operator when it is not desired.Finally, as the valve element 292 is moved to a nal position againstt-heopposing 'fluid pressure force of the throttle pressure, .passage 332 isbrought into fluid communication with the interior of valve element 292and with passage 414. This occurs when a valve port 42) is uncovered bythe associated valve lland on valve element 294. Simultaneously with theuncovering of port 420, valve element 292 blocks passage 380 so that thepressurized passage 382 will not be exhausted therethrough. After thevalve element 292 has assumed this ultimate position, the magnitude ofthe throttle pressure is equal to control pressure land this controlpressure is conducted to one side of the main shift valve spool throughpassage 382, thereby shifting the samein a left-hand direction to effecta downshift. The

release side of the brake servo 294 is exhausted through passage 372 andpassage 378 after a downshift occurs and since the apply side of theservo 264 continues to be pressurized, the brake band will be energized.

Concurrently with the application of brake band 169, clutch servo 210 isexhausted l'through passage 246,

through the upshifted hill hold valve 243 `and through passages 372 and37S. The blocker valveY 386 assumes a right-hand position during adownshift since the lefthand side of the blocker valve element ispressurized. Communication is thus established between control pressurepassage 254 and controltpassage 352 so that the latter s continues to bepressurized with control pressure.

Referring next to 4the adjustable pump exit feature previouslymentioned, the automatic controls therefor may be seen in FIGURE 3aandthey include a first performance valve 424 and a second performancevalve shown at 426. The servo for the adjustable blade elements 19,6V

of the pump exit section is pressurized when the blade elements are inthe normal cruising position and spring` means are provided forreturningthe blade elements to a high performance position when thepressure is released. The selective distribution of fluid pressure totheV pumpexit servo is controlled by lthe performance valves 424 and426.

Referring rst to valve 424, yit comprises asimple valvespool which isspring urgedv in an upward direction and which is urged in a downwarddirection by throttle pressure distributed thereto through the passage3M. When the valve assumes a downward position, communication isestablished between line pressure passage 24d and a passage 42Sextending to the servo for the pump exit section. The lower end of thevalve spool for performance valve 42d communicates with passage 380which is pressurized when the manual valve is shifted -to the HBposition. When the valve spool for performance valve 42d assumes anupward position under the influence of control pressure, communicationbetween passages 244 and 428 is established and the pump exit servo ispressurized. The pump may be conditioned for high performance byincreasing the engine throttle setting, which results in an increase inthrottle pressure and this in turn causes the erformance valve 424 toassume the high performance position.

The second performance valve 426 comprises a simple valve spool which isnormally urged in a downw -rd direction by converter pressure and whichis urged in an upward dhection by governor pressure distributed into thepassage 328. At relatively high vehicle speeds the governor pressure issufficiently high to move the valve spool for 4the second performancevalve 425. in an upward position and this causes an interruption betweenperformance valve 424 and passage 38d. Concurrently, however, controlpressure passage 25d communicates with the lower end of valve 424 sothat a downshift resulting from kick down pressure cannot beaccomplished at an undesirably high vehicle speed. The converterpressure passage 39? may be regulated by a simple converter pressureregulator valve identied by numeral 3%. Valve 431i comprises a simplevalve spool spring urged in an upward direction against the `fluidpressure force of lthe converter pressure acting on the upper endthereof. The balanced spring and pressure forces on the regulator valveelement control communication between the line pressure passage 432 andthe aforementioned converter pressure passage 3% to establish a reducedoperating pressure level in the hydrokinetic torus circuit of themechanism.

A lubrication pressure regulator valve is identified by numeral 433 andit comprises a simple valve spool 434 having spaced valve lands, one ofwhich is adapted to control the degree of communication of line pressurepassage 32 with a lubrication pressure passage 43d. The valve spool 434is urged in a left-hand direction as viewed in FIGURE 4a by a valvespring 438 and this spring force is opposed by the fluid pressure forceof the lube pressure in passage 436 which is distributed to theleft-hand side of valve spool 434. It is thus apparent that theregulator pressure in passage 43o will be reduced in magnitude and willbe determined by the calibration of the regulator valve 433.

A valve plunger 44MB is situated in the valve chamber for valve spool434 and the valve spring 38 is seated thereon as indicated. A passage442 communicates with the end of the valve chamber for regulator valve433 and it also communicates with the manual valve at a location whichis spaced from the valve lands 24%? and 242. Passage 442 alsocommunicates with reverse brake band servo 196 for the purpose ofdistributing control pressure to one side of the piston l98.

To obtain reverse drive operation, the manual Valve may be shifted tothe position indicated by the symbol R and this causes the passage 442to communicate directly with line pressure passage 244 through thespaced valve lands 249 and 242. Reverse brake servo 196 thereforebecomes pressurized and control pressure is also distributed to one sideof the plunger 440 of the regulator valve 442. This causes the valvespool 444 to move in a lefthand direction to bring passages 432 intocommunication with passage 436, thereby pressurizing the working chamberdened by the servo cylinder member illu and the annular piston 118 ofthe secondary turbine member. The annular piston 118 is thereforeshifted in a left-hand direction to rotate the blade elements 126 to areverse 22 position illustrated by means of full lines in FIGURE 6,thereby adapting the secondary turbine for reverse torque delivery.During normal forward driving operation, the blade elements 25 aremaintained -in a forward driving position illustrated in FlGURE 5 bydotted lines by reason of the torque exerted thereon by the torus flow.

When passage 442 is pressurized in this fashion, control pressure isdistributed to the left side of the aforementioned shuttle valve 252,thereby blocking communication between passages 256 and 254. Thisprevents fluid pressure in the reverse servo from being exhaustedthrough passage 256 which lbecomes uncovered by the valve land 242 ofthe manual valve when the manual valve spool is moved to a reverse driveposition.

Passage d2 also distributes control pressure to the lefthand side ofvalve plug 346 of the compensator valve and the spacer element 356 issubjected to control pressure. The control pressure force acting onspacer element 350 is transmitted directly to the compensator valvespool 3-32 to oppose the governor pressure force acting on the other endof the valve spool 332. The magnitude of the compensator pressurepassage M2 is thereby decreased by reason of this rearrangement offorces acting on compensator valve spool 332 and, as previouslyexplained, this results in an increase in the magnitude of the controlpressure. This increase in control pressure is necessary since thecapacity of the reverse servo must be suflicient to accommodate areverse driving torque.

ln addition to this variation in the servo capacity during reversedrive, the capacity of the low servo and the forward drive clutch servoduring forward drive operation may be varied in accordance with enginethrottle setting upon movement of fthe engine throttle from a closedposition to a setting of yapproximately 60% of the wide open position.The engine torque is generally proportional to throttle setting in thisrange of settings, but movement of the engine throttle beyond this rangeis not accompanied -by a signiiicant increase in eng-ine torque.'Provision is therefore made for making the throttle valve insensitiveto throttle movement at increased settings and this is done by adaptingthe throttle lvalve so that abutment 422 contacts shoulder 41S when thelimiting engine throttle setting is reached. Further increases in enginethrottle setting will therefore not affect throttle pressure. Thisfeature, together with the aforementioned governor pressure cutoutfeature in the compensator valve, insures a smooth shift under alldriving conditions.

We contemplate that variations in the preferred embodiment of ourinvention herein disclosed will come within the scope of our inventionas deiined by the following claims.

We claim:

l. In a power transmission mechanism, a driving member, a driven member,a gear unit having one portion thereof connected to said driven member,a hydrokinetic unit comprising a pump member and a turbine membersituated in toroidal fluid flow relationship, said pump member beingconnected to said driving member and said turbine member being connectedto another portion 0f said gear unit, said turbine member having flowdirecting blades disposed in spaced relationship about an axis ofrotation for said turbine member, said blades being arranged in twosections, one section being situated at the turbine iiuid flow entranceregion and the other being situated at the turbine fluid flow exitregion, means for adjustably positioning the blades of one sectionthereby altering the relative angular relationship `between the fluidiiow velocity vectors at entrance and exit regions of said turbinemember, said turbine member being adapted to deliver a forward drivingtorque to said driven member through said gear unit when said sectionsassume one relative angular relationship and to deliver a reversedriving torque to said driven member through said gear unit when saidsections assume another relative angular relationship.

2. In a power transmission mechanism, a driving mem- 23 ber, a drivenmember, agear unit comprising a sun gear element, a ring gear element, acarrier element and planet gears carried by said carrierV element inmeshing engagement with said sun and ring gear elements, means fordrivablyconnecting one element of said gear unirt to said driven member,brake means for anchoring a second element of said gear unit torprovidea torque reaction, a hydrokinetic unit comprising a pump member, aprimary turbine member and a secondary turbine member, said pump andturbine members being disposed in uid fiow relationship, said pumpmember being connected to said drivingmember, means for transmitting thetorque of each ofsaid turbines to said gear -unit during for-Ward driveoperation and brake means -for anchoring one of said turbine members toobtain reverse drive operation whereby a reverse driving torque isimparted to the other turbine member and transmitted to an element ofSaid gear unit.

3. In a power transmission mechanism, a gear unit Y comprising a sungear member, a ring gear member, a

-carriervmember and planet gears mounted on said carrier member inmeshing engagement with said sun and ring -gear members; a first memberof said gear unitfbeing connected to driven portions of said mechanism,the hydrokinetic unit comprising a pumpmernber, a primary'turbine memberand a secondary turbine member situated in uid ow relationship, meansforv transmitting the torque of each turbine member through said gearunit to said driven portions during forward drive operation, brake meansfor anchoring a second member of said gear unit to absorb the forwarddriving torque reaction, and additional means for braking one turbinemember and one member ofv said gear unit to obtain reverse driveoperation, the torque acting on the other turbine member being reversedin direction when said additional brake means is actuated, said gearunit being effective to transmit the reverse driving torque of saidother turbine member to said driven portions during reverse driveoperation.

4. In a power transmission mechanism, a gear unit comprising a sun gearymember, a ring gear member, a carrier member and planet gears mounted onsaid carrier member in meshingengagement with said sun and ring gearmembers, a rst member of said gear-unit being Vconnected to drivenportions of said mechanism, a hydrokinetic unit comprising a pumpmember, a primaryV turbine member and a secondary turbine membersituated Vin fluid ow relationship, means for transmitting the ber and athird-member of said gear unit to condition said mechanism for reversedrive operation, the torque acting on the secondary turbine member beingreversed in direction whenrsaid additional brake means is actuated, saidgear unit being effective to transmit the. reverse drivingV torque ofsaid turbine member to said driven portions during reverse driveoperation.

5. In a power transmission mechanism, a gear unit comprising a sun gearmember, a ring gear member, a carrier member and planet gears mounted onsaid car-` rier member in meshing engagement with said sun and ring gearmembers, Vaiirst member of said gear unit being connected todrivenportions of said gear unit, a hydrokinetic unit comprising a pumpmember, a primary 'turbine member and a secondary turbine member.situbrake means for anchoring a third member of said gear unit toabsorb the forward driving torque reaction, and additional brake meansfor braking said primary turbine member and said second member of saidsecond gear unit to obtain reverse dr-ive, clutch means for connectingsaid secondary turbine member to said third member of said gear unitduring reverse drive, the torque acting on the secondary turbine memberbeing reversed in direction when said additional brake means isactuated, said gear unit being effective to transmit the reverse drivingtorque of said secondary turbine member to said driven portions.

6. In a power transmission mechanism, a driving member, a driven member,a gear unit comprising a sun gear element, a ring gear element, acarrier element and planet gears mounted on said carrier element inmeshing engagement with said sun and ring gear elements, means fordrivably connecting one element of said gear unit to said driven member,brake means forranchoring a second element ofsaid gear unit to provide atorque reaction, a hydrokinetic unit comprising a pump member, a primaryturbine member and a secondary turbine member, said pump and turbinemembers being disposed in uid ow relationship, said pump member beingconnected to said driving member, means for connecting saidV primaryturbine member to an element of said gear unit, overrunning clutch meansfor forming a one-way driving connection between said secondary turbinemember and said primaryY bine member to another element of said gearunit whensaid mechanism is conditioned for reverse drive operationwhereby a reverse driving torque is imparted to said secondary turbinemember and transmitted to said driven member through said gear unit.

7. In a power transmission mechanism, a driving member, fa drivenmember, a gear unit comprising a sun gear member, a ring gear member, acarrier member and planet gears mounted on said carrier member inmeshing engagement with said sun and ring gear members, a rst memberV ofsaid gear unit being connected tovsaid driven member, brake means foranchoring a'second member of said gear unit to provide a forward drivingtorque `reaction, a hydrokinetic unit comprising a pump member connectedto said driving member, a primary turbine member Iand a secondaryturbine member, the members of said hydrokinetic unit being disposedVin'uid ow relationship, the primary member being drivably connectedtor-a third member of said gear unit, overrunning clutch means nforconnecting said secondary turbine member to said third member of saidgear unit, said overrunning clutch means for-ming a driving torquedelivery path during forward drive operation and permitting rotation ofsaid secondary turbine member in a direction opposite to the directionof rotation of said pump member during reverse drive operation,additional brake means for anchoring said primary turbine member andsaid third member of said gear unit and clutch means for coupling saidsecondary turbine member to said second member of said gear unit tocondition said mechanism for reverse operation.

8. VIn a power transmission mechanism, -a driving member, a drivenmember, a gear unit comprising a sun gear member, a ring gear member, acarrier member and planetgears mounted on said carrier member .inmeshing engagement with said sun and ring gear members, the carriermember of said gear unit being connectedY to said first `gear member,brake means for anchoring the sun gear member of said gear unit toprovide a torque reaction, a hydrokinetic unit comprising a pump memberconnected to said driving member, a primary turbine member and asecondary turbine member, the members of said hydrokinetic unit beingdisposed in uid flow relationship, the

e primary turbine member being directly connected to the ring gearmember of said gear unit, overrunm'ng clutch means'for connecting saidsecondary turbine member to said ring gear member, said overrunningclutch means forming an overrunning torque delivery path during forwarddrive operation and permitting rotation of said secondary turbine memberina direction opposite `to the direction of rotation of said pump memberduring reverse drive operation, additional brake'means for anchoringsaid primary turbine land -said ring gear and clutch means for couplingsaid secondary turbine member to said sun gear to condition `saidmechanism for reverse drive operation.

9. In a power transmission mechanism, a driving member, a driven member,a gear unit comprising a sun gear member, a ring gear member, a carriermember and planet gears mounted on said carrier member in meshingengagement with said sun and ring gear members, the carrier member ofsaid gear unit being connected to said driven member, brake means foranchoring the sun gear member of said gear unit to provide a torquereaction, a hydrokinetic unit comprising a pump member anchored to saiddriving member, a primary turbine member and a secondary turbine member,the members of said hydrolrinetic unit being disposed in fluid owrelationship, the primary turbine member being directily connected tothe ring gear member of said gear unit, overrunning clutch means forconnecting said secondary turbine member to said ring gear member, saidoverrunning clutch means forming a one-way torque delivery path duringforward drive operation and permitting rotation of said secondaryturbine member in a direction opposite to the direction of rotation ofsaid pump member during reverse drive operation, and additional brakemeans for anchoring said primary turbine and said ring gear member andclutch means for coupling said secondary turbine member to said sun gearmember to condition said mechanism for reverse drive operation, saidsecondary turbine member being formed with a first blade sectiondisposed at the fluid entrance region thereby and a second blade sectiondisposed at the fluid flow exit region thereof, each section includingflow directing blades, the blades of the second section being adjustablerelative to the blades of the first section to provide a variation inthe eective secondary turbine exit angle, the blades of said secondsection assuming one position during forward drive whereby aV forwarddriving torque is imparted to the secondary turbine member and assuminga second position during reverse drive operation whereby a reversesecondary turbine torque is produced.

10. In a power transmission mechanism, a gear until comprising aV sungear member, a ring gear member, a carrier member and planet gearsmounted on said carcarier member in meshing engagement with said sun andring gear members, a tirst member of said gear unit being connected todriven portions of said mechanism, a hydrokinetic torque convertercomprising a pump member, primary and secondary turbine members and areactor member, said pump member being connected to a driving portion ofthe mechanism, means for simultaneously transmitting the torque of eachturbine member of said hydrokinetic unit to a second member of said gearunit during forward drive operation, brake means for anchoring a thirdmember of said gear unit to absorb the forward. driving torque reaction,and additional brake means for braking said second member of said gearunit and clutch means for connecting a turbine member of saidhydrokinetic'unit to said third member of said gear unit to conditionsaid mechanism for reverse drive operation, said gear unit beingeffective to transmit a reverse driving torque from said third member ofsaid hydrokinetic unit to said driven portions of said mechanism duringreverse drive operation.

11. In a power transmission mechanism, a gear unit comprising a sun gearmember, a ring gear member, a carrier member and planet gears mounted onsaid carrier member in meshing engagement with the sun and ring gearmembers, the carrier member being in driving relationship with respectto power output portions of the mechanism, brake means Afor selectivelyanchoring said sun gear member, a hydrokinetic unit comprising a, pumpmember, a primary turbine member and a secondary turbine member situatedin iluid flow relationship, each of the members of said hydrokineticunit comprising an inner and outer shroud defining in part the toroidalfluid now path, ow directing blade elements situated between the shroudsof each member, means for drivably connecting the primary turbine memberto said ring gear member, overrunning clutch means forming a one-waytorque delivery path between said secondary turbine member and said ringgear member whereby the driving torque of said secondary turbine membersupplements the driving torque of said primary turbine member, clutchmeans for selectively coupling said secondary turbine to said sun gearmember, reverse brake means for anchoring said primary turbine memberand said ring gear member including a friction member, a torque transfermember connected to said primary turbine member in the region of theinner shroud thereof andv extending in the direction of the geometricartis of the hydrokinetic unit between the secondary turbine member saidpump member, and a torque delivery shaft connecting said torque transfermember to said friction member, said mechanism being conditioned byreverse operation when said reverse brake and said clutch means areconcurrently actuated and when said first named brake means is released.

12. The combination as set -forth in claim 11 wherein said secondaryturbine member includes two bladed sections, one section being disposedat the fluid flow entrance region of the secondary turbine member andthe other being situated at the uid flow exit region thereof, the bladesof the other section of said secondary turbine being adjustable relativeto the blades of the rst section thereof to provide a change in thedirection of the effective torque acting on the secondary turbine.

13. In a hydrokinetic unit for use with a power transmission mechanismhaving a driving member and a driven member, a pump member and a turbinemember situated in duid ow relationship, said turbine member comprisinga bladed iiuid entrance section and a bladed fluid exit section, eachsection including flow directing blades, said pump member beingconnected to said driving member, means for connecting the turbinemember to said driven member, and means for adjusting the angularity ofthe blades of the exit section relative to the blades of the entrancesection to condition the turbine member for torque delivery in either aforward or reverse direction, said turbine member comprising a hubportion and said adjusting means comprising a Huid pressure operatedservo disposed within and defined in part by said hub portion, saidservo including a fluid pressure operated member and a mechanicalconnection between said fluid pressure operated member andthe blades ofthe exit section of said turbine member.

14. In a power transmission mechanism, a gear unit comprising a sun gearmember, a ring gear member, a carrier member and planet gears mounted 0nsaid carrier member in meshing engagement with said sun and ring gearmembers, means for braking a first member of said gear unit to provide atorque reaction, a second member of said gear unit being connected topower output portions of said mechanism, a hydrokinetic unit cornprisinga pump member, a primary turbine member and a secondary turbine member,the primary turbine member being connected to a third member of saidgear unit and one-way coupling means forming a driving con nectionbetween said secondary turbine member and said third member of said gearunit to supplement the driving torque of said primary turbine.

15. in a power transmission mechanism, a gear unit comprising a sun gearmember, a ring gear member, a

Y Y 27 carrier member and planet gears mounted on said carrier member inmeshing engagement with said sun and ring gear members, means forbraking a rst member of said gear unit to provide a torque reaction, asecond member of said gear unit being connected to power output portionsof said mechanism, the hydrokinetic unit comprising a pump member, aprimary turbine member and a secondary turbine member, the primaryturbine member being con nected to a third memberof said gear unit,one-way coupling means forming a driving connection between saidsecondary turbine member and said third member of said gear unit tosupplement the driving torque of said primary turbine member and meansfor preventing free 'reverse rotation of said secondary turbine memberduring forward driving operation.

16. The combination as set forth in claim wherein said last mentionedmeans includes a selectively operable clutch mechanism adapted toconnect said secondary turbine member to said Abraking means foranchoring said secondary turbine member during a part of the forwarddriving range.

17. In a power transmission mechanism, a gear unit comprising a sun gearmember, a ring gear member, a carrier member and planet gears mounted onsaid carrier in meshing engagement with the sun and ring gear members,means for braking a rst member of said gear unit to provide a torquereaction, a second member of saidk gear unit being connected to a poweroutput portion of said mechanism, a hydrokinetic unit comprising a pumpi ember, a primary turbine member and a secondary turbine member, themembers of the hydrokinetic unit being Ysituated in iiuid iiowrelationship, the primary turbine member being connected to one memberof said gear unit, overrunning coupling means forming a one-way drivingconnection between said secondary turbine kmember and said one member ofsaid gear unit to supplement the driving torque of the primary turbine,s id secondary turbine member comprising separate sections at the iluidow entrance and exit regions thereof, each section lcomprising flowdirecting blades, means for adjustably positioning the blades of theexit section relative to the blades of the entrance section, saidsecondary turbine member being adapted to deliver a negative torque tosaid gear unit when the blades of the exit section thereof assume onerelative angular position and to deliver a forward driving torque tosaid gear unit when the blades of the exit section thereof assumeanother relative angular position, additional brake means forseiectively anchoring said primary turbine member and said one member ofsaid gear unit and clutch means for selectively coupling said secondaryturbine member to said rst member of said gear unit, said iirst namedbraking means being deenergized and said last named brake means and saidclutch means being energized for conditioning the mechanism lfor reverseoperation.

18. rhe combination as set forth in claim 17 wherein said secondaryturbine member includes a hub portion and a servo mechanism situated inand partly defined by said hub portion, said servo mechanism including ailuid pressure operated member and a mechanical connection between saidfiuid pressure operated member and the blades of. the exit section ofsaid secondary turbine member for adjustably positioning the latter.

19.v in a power transmission mechanism for delivering power from adriving member to a driven'member, a hydrolcinetic unit comprisingprimary and secondary turbine members and a pump member disposed in uidflow relationship, said pump member being connected to said drivingmember, said primary turbine member being in driving relationship withrespect to said drivenA member and adapted to deliver a positive drivingtorque to said driven member during forward drive operation, and meansfor conditioning said secondary turbine member for reverse torquedelivery to eect reverse drive operation including a selectivelyengageable brake connected to said primary turbine member, said brakebeing adapted to anchor said primary turbine member when energized, andselectively engageable clutch means forming in part a reverse torquedelivery path between said secondary turbine member and said drivenmember, said clutch means and said brake being energized during reversedrive.

2G, In a power transmission mechanism for delivering power from ad-riving member to a driven member, a hydrokinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed iniiuid ilow relationship, said pump member being connected to saiddriving member, a planetary gear unit forming in part separate torquedelivery paths between said driven member and each bladed torquetransmitting member, a iirst of said bladed torque transmitting membersbeing connected directly to a irst power input gear member of said gearunit whereby forward driving torque is delivered to said gear member'during forward drive operation, a power output member of said gear unitbeing connected to said driven member, means including a selectivelyengageable clutch for connecting the other bladed torque transmittingmember to another power input member of said gear unit during reversedrive operation, first brake means for anchoring selectively said iirstbladed torque transmitting member and said first power input membervpower input member functioning as a reaction member for said gear unitduring forward drive operation.

2l, in a power transmission mechanism for delivering power from adriving member to a driven member, a hydrolrinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed iniiuid iow relationship, said pump member being connected to said drivingmember, a planetary gear unit forming in part separate torque deliverypaths between said driving member and each bladed torque transmittingmember, said gear unit comprising a ring gear member, a sun gear member,a carrier member and planet gears carried by said carrier lmember inmeshing engagement with said `sun and ring gear members, a iirst of saidbladed torque transmitting members being connected directly to said ringgear member of said gear unit whereby forward driving torque isdelivered to said gear unit during forward drive operation, the carriermember of said gear unit being connected to said driven member, Ymeansincluding a selectively engageable clutch for connecting the otherbladed torque transmitting member to said sun gear member of said gearunit during reverse drive operation, iirst brake means for anchoringselectively said first bladed torque transmitting member and said ringgear member of said gear unit during reverse drive operation, and secondbrake means for anchoring said sun gear during forward drive operation,said ring gear functioning as a reaction member for said gear unitduring reverse drive operation and said sun `gear functioning as areaction member for said gear unit during forward drive operation. Y

22. In a power transmission mechanism for delivering power from adriving member to a driven member, a hydroltinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed influid ilow relationship, said pump member being connected to saiddriving member, a planetary gear unit forming in part torque deliverypaths between said driven mem-ber and each bladed torque transmittingmember, a rst of said bladed torque transmitting members being connectedVdi-V rectly to a first power input gear member of said gear unitwhereby forward driving torque is delivered to said gear member duringforward drive operation, a power output member of said gear unit beingconnected to said driven member, means for Yconnecting the other bladedtorque transmitting member to another power input member of said gearunit during reverse drive operation, first brake means for anchoringselectively said first bladed torque transmitting member and said firstpower input member during reverse drive operation, second brake meansfor anchoring said second power input member of said gear unit duringforward drive operation, said first power input member functioning as areaction member for said gear unit during reverse drive operation andsaid second power input member functioning as a reaction member for saidgear unit during forward drive operation, said second bladed torquetransmitting member being subjected to a hydrokinetic torque reactionduring forward drive operation that is opposite in direction to thehydrokinetic torque reaction acting upon said first blade torquetransmitting member, and coupling means for transferring said oppositelydirected torque reaction of said second bladed torque transmittingmember to said second brake means to inhibit rotation of said secondbladed torque transmitting member in said opposite direction.

23. In a power transmission mechanism for delivering power from adriving member to a driven member, a hydrokinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed influid fiow relationship, said pump member being connected to saiddriving member, a planetary gear unit forming in part separate torquedelivery paths between said driving member and eacli bladed torquetransmitting member, said gear unit comprising a ring gear member, a sungear member, a carrier member and planet gears carried by said carriermember in meshing engagement with said sun and ring gear members, afirst of said bladed torque transmitting members being connecteddirectly to said ring gear member of said gear unit whereby forwarddriving torque is delivered to said gear unit during forward driveoperation, the carrier member of said gear unit being connected to saiddriven member, means including a selectively engageable clutch forconnecting the other bladed torque transmitting member 'to said sun gearmember of said gear unit during reverse drive operation, first brakemeans for anchoring selectively said first bladed torque transmittingmember and said ring gear member of said gear unit during reverse driveoperation, second brake means for anchoring said sun gear during forwarddrive operation, said ring gear functioning as a reaction member forsaid gear unit during reverse drive operation and said sun gearfunctioning as a reaction member for said gear unit during forward driveoperation, said second bladed torque transmitting member being subjectedto a hydrokinetic torque reaction during forward drive operation that isopposite in direction to the direction of the hydrokinetic torquereaction acting upon said first bladed torque transmitting member, andcoupling means for transferring said oppositely directed torque reactionof said second bladed torque transmitting member to said second brakemeans to inhibit rotation of said second bladed torque transmittingmember in said opposite direction.

24. In a power transmission mechanism for delivering power from adriving member to a driven member, a hydrokinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed influid ow relationship, said pump member being connected to said drivingmember, a planetary gear unit forming in part torque delivery pathsbetween said driven member and each bladed torque transmitting member, arst -of said bladed torque transmitting members being connected directlyto a first power input gear member of said gear unit whereby a forwarddriving torque is delivered to said gear member during forward driveoperation, a power output member of said gear unit being connected tosaid driven member, means for connecting the other bladed torquetransmitting member yto another power input member of said gear unitduring reverse drive operation, first brake means for anchoringselectively said first bladed torque transmitting member and said firstpower input member during reverse drive operation and second ibrakemeans for anchoring said second power input member of said gear unitduring forward drive operation in `a first forward driving speed ratio,said rst power input member functioning as a reaction member for saidgear unit during reverse drive operation and said second power inputmember functioning as a reaction member for said gear .unit duringforward drive operation in said first forward driving speed ratio, andselectively engageable direct drive clutch means for coupling togetherthe members -of said gear unit for conjoint rotation to establish asecond high yspeed forward driving speed ratio.

25. In a power transmission mechanism for delivering power from adriving member to a driven member, a hydrokinetic unit comprising twobladed torque transmitting members and a bladed pump member disposed influid flow relationship, said pump member being connected to saiddriving member, a planetary gear unit forming in part separate torquedelivery paths between said driving member and each bladed torquetransmitting member, said gear unit comprising a ring gear member, a sungear member, a carrier member and planet gears carried by said carriermember in meshing engagement with said sun and ring gear members, afirst of said bladed torque transmitting members being connecteddirectly to said ring gear member of said gear unit whereby forwarddriving torque is delivered to said gear unit during forward driveoperation, the carrier member of said gear unit being connected to saiddriven member, means for connecting the other bladed torque transmittingmember to said sun gear member of said gear unit during reverse driveoperation, first brake means for anchoring selectively said first bladedtorque transmitting member and said ring gear member during reversedrive operation, second brake means for anchoring said sun gear duringforward drive operation in a first forward driving ratio, said ring gearfunctioning as a reaction member for said gear unit during reverse driveoperation and said sun gear functioning as a reaction member for saidgear unit during forward drive operation in said first forward drivingspeed ratio, and selectively engageable direct drive clutch means forcoupling together the members of said gear unit for conjoint rotation toestablish a second high speed forward driving speed ratio.

References Cited in the file of this patent UNITED STATES PATENTS1,298,990 Mason Apr. l, 1919 2,205,794 Jandasek June 25, 1940 2,292,482lRoche Aug. 11, 1942 2,293,358 Pollard Aug. 18, 1942 2,578,450 PollardDec. 11, 1951 2,695,533 Pollard Nov. 30, 1954 2,719,616 Ahlen Oct. 4,1955

1. IN A POWER TRANSMISSION MECHANISM, A DRIVING MEMBER, A DRIVEN MEMBER,A GEAR UNIT HAVING ONE PORTION THEREOF CONNECTED TO SAID DRIVEN MEMBER,A HYDROKINETIC UNIT COMPRISING A PUMP MEMBER AND A TURBINE MEMBERSITUATED IN TOROIDAL FLUID FLOW RELATIONSHIP, SAID PUMP MEMBER BEINGCONNECTED TO SAID DRIVING MEMBER AND SAID TURBINE MEMBER BEING CONNECTEDTO ANOTHER PORTION OF SAID GEAR UNIT, SAID TURBINE MEMBER HAVING FLOWDIRECTING BLADES DISPOSED IN SPACED RELATIONSHIP ABOUT AN AXIS OFROTATION FOR SAID TURBINE MEMBER, SAID BLADES BEING ARRANGED IN TWOSECTIONS, ONE SECTION BEING SITUATED AT THE TURBINE FLUID FLOW ENTRANCEREGION AND THE OTHER BEING SITUATED AT THE TURBINE FLUID FLOW EXITREGION, MEANS FOR ADJUSTABLY POSITIONING THE BLADES OF ONE SECTIONTHEREBY ALTERING THE RELATIVE ANGULAR RELATIONSHIP BETWEEN THE FLUIDFLOW VELOCITY VECTORS AT ENTRANCE AND EXIT REGIONS OF